KR940009227B1 - Method of operating heat pump - Google Patents

Abstract

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Description

Heat pump system

1A and 1B are schematic circuit diagrams of a basic two stage heat pump and a first stage heat pump to which the present invention is applied, respectively.

2A, 2B and 2C are a plan view, a side view, and a partially enlarged view of one embodiment of a condenser for a heat pump system of the present invention, respectively.

3A and 3B are performance diagrams of a two-stage heat pump according to the present invention, where FIG. 3A is a temperature gradient diagram of one embodiment of a refrigerant, and FIG. 3B is a mollier diagram.

Figures 4a and 4b is a performance diagram of another embodiment of a heat pump according to the present invention, Figure 4a is a temperature gradient diagram of the heat pump cycle through another embodiment of the refrigerant, Figure 4b is a Moliere diagram.

5A and 5B are performance diagrams of a conventional heat pump operation based on the two-stage heat pump unit of FIG. 1A, where FIG. 5A is a temperature gradient diagram and FIG. 5B is a Moliere diagram.

* Explanation of symbols for main parts of the drawings

1,11 compressor 2: condenser

3,13: expansion valve 4,14: evaporator

5,15: heat storage 22: series condenser / evaporator

30: double tube 31: outer tube

32: inner tube 33: wire fin

The present invention relates to a heat pump system comprising a condenser which allows the discharge of high quality fluids, ie extremely hot fluids (eg water) and a large temperature difference between the inlets and outlets.

In particular, the present invention relates to a heat pump system characterized by increasing the coefficient of performance (COP) and the heat exchange amount by increasing the subcooling of the refrigerant condensate in the condenser.

Conventionally, various heat pumps for discharging high temperature fluids, low temperature fluids or high and low temperature fluids are known.

Generally, a two stage heat pump consists of a high temperature side heat pump cycle for a high boiling point refrigerant and a low temperature side heat pump cycle for a low boiling point refrigerant as shown in FIG. 1b, and the high temperature heat output is discharged from a high temperature side heat pump cycle. Low temperature heat output is released from the low temperature side heat pump cycle. Examples of such two-stage heat pumps are Japanese Patent Laid-Open Nos. 62-52376 (1987) and 62-52377 (1987), Japanese Utility Model Publication Nos. 57-2364 (1982) and 63-2053. 1988, et al. Heat pumps comprising a single stage heat pump as shown in FIG. 1b are also known.

In such heat pump units, the coefficient of performance and the temperature at which the heated fluid is obtained usually varies with the following factors for a given refrigerant: (a) the performance of devices such as compressors, condensers, evaporators, etc., (b) condensation The state of the two fluids undergoing mutual heat exchange over time, i.e. the condensation temperature of the refrigerant, the condenser inlet temperature and the flow rate of the fluid to be heated (e.g. water), (c) the evaporation state (flow rates of the two fluids undergoing mutual heat exchange on the evaporator side, Temperature) etc.

The performance coefficient of the heat pump and the acquisition temperature of the fluid to be heated are determined by the heat exchange state on the condenser side, and the evaporation state (flow rate and temperature) becomes clear when the refrigerant (refrigerants), the compressor (compressors) and the evaporator are preset. .

The conventional two stage heat pump is operated according to a refrigeration cycle which shows a temperature gradient diagram (heat exchange state between the refrigerant and the fluid in the condenser) as shown in FIG. 5A, and the superheated gas at the condenser inlet (the outlet side of the compressor). The refrigerant on the hot side maintained in state “a” undergoes a temperature change to reach “b” (saturated steam line), and changes from saturated steam state “b” to saturated liquid state “c” in the condenser (b to c). And from the condenser outlet (inlet of the expansion valve) to the subcooled liquid state “d” (c to d).

On the other hand, the aforementioned Moliere diagram shows in Figure 5b that the refrigerant is compressed in the compressor (from f to a) and the enthalpy changes from "i ' ₆ " to "i' ₁ ": The refrigerant vapor passes through the saturated steam line “b” and the enthalpy changes from “i ′ ₁ ” to “i ′ ₂ ”, is cooled by heat exchange with the liquid to be heated in the condenser and liquefied under constant pressure (b to c) During this process, the enthalpy “i ′ ₂ ” is changed from “i ′ ₂ ”: After passing through the point “c” (i ′ ₃ ) of the saturated liquid line, the refrigerant liquid is subcooled (c to d) ₄ 'state.

In FIGS. 5A and 5B, the symbols “t′w ¹ ” and “t′w ² ” represent the condenser inlet temperature and the condenser outlet temperature of the liquid to be heated, and the symbols “e” and “f” indicate the in series condenser. The inlet and outlet states are shown.

In this way, conventional refrigeration cycles are operated to ensure some degree of subcooling in order to operate without damaging the expansion valve, and at present, the degree of subcooling required to operate the expansion valve is generally considered to be about 3 ° C to 5 ° C.

In this respect, the conventional heat pump and its condensers have a maximum heat transfer coefficient in the saturation region of the refrigerant, but the heat transfer characteristics in the superheated and supercooled regions of the refrigerant are not so high, so the condenser becomes large and reduces the economic advantages. Let's do it. In addition, it is thought that a method of increasing the supercooling degree to avoid the decrease of the coefficient of performance due to the increase in the pressure loss caused by the enlargement of the condenser is not recommended.

In conventional heat pumps, non-reverse flow type heat exchangers such as shell-tube heat exchangers and heat exchangers of bottle flow type are adopted. The problem with these heat exchangers is that the coolant liquid is not sufficiently cooled because the lowest inlet temperature of the water to be heated cannot be obtained. For example, a heat exchanger of a traveling flow type cannot lower the temperature of the refrigerant liquid below 90 ° C when the water temperature at the outlet is 90 ° C. Even in the case of a reverse flow type heat exchanger, if a procedure for gradually raising the temperature of the heated water while circulating the water is taken, the inlet temperature of the water for cooling the refrigerant liquid rises over time to a high temperature. . For example, if the inlet temperature of water is 90 ° C in that case, the refrigerant liquid cannot be made lower than 90 ° C. When the flow rate of the circulating water becomes large while the heat exchange amount is maintained, the temperature difference of the water between the outlet temperature and the inlet temperature must be small. As a result, there is still a problem that the inlet temperature is too high for the same outlet temperature so that the refrigerant liquid cannot be sufficiently cooled.

In view of the various problems mentioned above with conventional heat pumps, the present invention is particularly intended to solve the problems by considering the condenser design and its performance conditions.

Accordingly, it is a main object of the present invention to provide a heat pump system that provides an increased coefficient of performance.

Another object of the present invention is to provide a heat pump system capable of producing high quality energy, ie hot steam or hot water that can be used. The present invention, which is suitable for the above-mentioned objects, is based on satisfying the following requirements in order to obtain the outlet temperature of hot water and to obtain a large degree of subcooling.

(a) Adopt a full reverse flow heat exchange system. (b) The condenser inlet temperature of the heated fluid (water) is made lower than the desired outlet temperature of the refrigerant liquid from the condenser. (c) make the flow rate of the fluid to be heated relatively small, as long as there is no inconvenience in discharging.

That is, according to one aspect of the invention, there is provided a heat pump system comprising a heat pump cycle for a refrigerant consisting of a compressor, a condenser, an expansion valve and an evaporator, the condenser includes a through passage for the heated fluid, The condenser consists of a heat exchanger in the form of a complete reverse flow between the refrigerant and the fluid, and when the system is operated, the refrigerant is liquefied and condensed while the refrigerant transfers heat to the fluid in a reverse flow manner, followed by a saturation temperature of the refrigerant. The condenser is operated such that it is subcooled with a subcooling of at least 20% of the temperature difference between the condenser inlet temperature of the fluid and the hot fluid is discharged from the outlet of the condenser.

In the heat pump system described above, the evaporator further comprises a through passage for the fluid to be cooled and becomes a heat exchanger of a completely reverse flow type between the refrigerant and the fluid to be cooled, and only the low temperature fluid is discharged or both the high temperature and the low temperature fluid are discharged. do.

According to another aspect of the present invention, there is provided a high temperature side heat pump cycle for a high boiling point refrigerant including a compressor, a condenser, an expansion valve, and a heat accumulator coupled by a pipe in sequence; A low temperature side heat pump cycle for low boiling point refrigerant comprising an extruder, an expansion valve, an evaporator and a heat accumulator coupled in sequence by a pipe; There is provided a two stage heat pump system consisting of a series condenser / evaporator which mutually couples the condenser of the hot cycle and the evaporator of the cold cycle in a heat exchangeable manner; The condenser further comprises a through passage for the fluid to be heated, wherein the condenser is configured as a heat exchanger of a complete reverse flow type between the high boiling point refrigerant and the fluid to be heated, and the high boiling point refrigerant is to be heated when the system is operated. During the heat exchange in a reverse flow manner with the fluid, the high boiling point refrigerant is liquefied and condensed, and then the condenser is operated to subcool with a subcooling of at least 20% of the temperature difference between the saturation temperature of the refrigerant and the condenser inlet temperature of the fluid to be heated. , Hot fluid is discharged.

In the two stage heat pump system, the evaporator further includes a through passage for the cooled fluid, and the hot and cold fluids are discharged.

In the case of the two heat pump units described above, the condenser is preferably formed in a double or multi-pipe comprising an outer tube and an outer corrugated tube with wire fins, the fluid to be heated is moved through the inner tube and the refrigerant is It is transferred in a reverse flow manner to the fluid to be heated through the space between the inner and outer tubes.

Throughout this specification and claims, the term “supercooling” refers to the temperature difference between the refrigerant saturation temperature and the condenser outlet temperature of the refrigerant liquid.

According to the present invention, when the heat pump system is operated, the refrigerant is evaporated in the evaporator and sucked into the compressor to be discharged as a gas at a high temperature and high pressure, and liquefied into a condensed liquid in the condenser to provide heat to a fluid to be heated which flows back into the refrigerant. After passing through the expansion valve, the gas-liquid mixture is returned to the evaporator. For example, in a condenser where the refrigerant is a high boiling point S-dichlorotetrafluoroethane, the refrigerant is supercooled with a subcooling of at least 18.6 ° C because its saturation temperature is about 112 ° C, and the inlet temperature of the fluid to be heated is brought to normal temperature. Assuming about 19.1 ° C, the subcooling is calculated as "(112-19.1) x 0.2 = 18.6 (° C). This value is considerably higher than conventional subcoolings of 3 ° C to 5 ° C. In general, the higher the degree of subcooling, the larger the coefficient of performance, and thus the present invention clearly improves the coefficient of performance.

Furthermore, since the enthalpy difference of the refrigerant between the saturation state in the condenser and the supercooling state at the outlet of the condenser is larger than the enthalpy difference with respect to the conventional heat pump, the temperature difference between the inlet and outlet temperatures of the fluid to be heated is the temperature difference of the conventional heat pump. Greater than As a result, hot steam (about 120 ° C.) or hot fluid (about 100 ° C.) is discharged with high efficiency. It is also possible for the fluid to have a temperature higher than the saturation temperature of the refrigerant in the condenser.

Refrigerant (s) used in the heat pumps of the present invention are, for example, 1,1,2-trichloro-1,2, 2-trifluoroacetane (Flon R-113), S-dichloro tetrafluorouroethane Fluorohydrocarbons such as (Flon R-114), dichloro difluoromethane (Flon R-12), chlorodifluoromethane (Flon R-22) and the like.

With a two-stage heat pump system, high-boiling refrigerants such as Flon R-113, Flon R-114, and Flon R-11 are used for high temperature cycles, while low-boiling refrigerants such as Flon R-12, Flon R-22, etc. Used for cycles.

EXAMPLE

The invention is described with reference to the accompanying drawings.

The invention is applicable to a two stage heat pump unit or a one stage heat pump unit depending on the purpose of the fluid to be discharged.

The two-stage heat pump shown in FIG. 1a typically comprises a compressor 1 which is coupled in sequence, a condenser 2 having a through passage for the fluid to be heated, an expansion valve 3 and a regenerator 5. Hot side cycles for point refrigerants; It consists of a low temperature side cycle for low boiling point refrigerant comprising an compressor 11, an expansion valve 13, an evaporator 14 having a through passage for the fluid to be cooled, and a heat accumulator 15, which are coupled in sequence, The condenser 2 of and the evaporator 14 of the low temperature side cycle are coupled in a heat exchange manner via a series condenser / evaporator 22.

On the other hand, the first stage heat pump unit has a compressor (1), a condenser (2) having a through passage for the fluid to be heated, an expansion valve (3), a fluid to be cooled, as shown in FIG. It consists of an evaporator (4) and a heat storage (5) having a through passage for the.

In both cases, the fluid to be heated introduced from the fluid source into the inlet 6 of the condenser 2 is transferred to the refrigerant through the condenser in full reverse flow to exchange heat with the refrigerant, and the hot fluid or vapor It is necessary to discharge from the outlet 7 of the condenser 2 in the state.

The fluid to be cooled and the refrigerant to be supplied from the inlet "8" or "18" of the evaporator "4" or "14" preferably flow in a reverse flow through the evaporator "4" or "14". In the case of a two stage heat pump unit, the high boiling point refrigerant and the low boiling point refrigerant are preferably flowed through the series condenser / evaporator 22 in a reverse flow manner.

3a and 3b respectively show a temperature gradient diagram and a molier diagram of a two-stage heat pump system according to the present invention, wherein the high boiling point refrigerant is S-dichlorotetrafluoroethane (Flon R-114), respectively. Compared with the conventional heat pump operation of FIGS. 5A and 5B.

As is evident in the temperature gradient comparison (Figs. 3a and 5a), the high boiling point refrigerant maintained in superheated gas state "A" at the inlet of the condenser 20 according to the invention is saturated liquid in saturated gas state in the condenser. State (B to C), and the state change is similar to the conventional operation method shown in FIG. 5A. However, the significant change is that the refrigerant in saturated liquid state changes from the outlet of the condenser 2 to the subcooled state “D” and shows a large degree of subcooling, on the water side between the outlet t _w2 and the inlet t _{w1 of the} condenser. It can be seen that the temperature difference is large compared with the conventional operation.

In the Moliere diagram of FIG. 3b of the heat pump cycle according to the present invention, in the process of the refrigerant in superheat state “A” at the inlet 6 of the condenser 2 becoming saturated gas state “B”, the enthalpy is “ from i ₁ ”to“ i ₂ ”; The enthalpy changes to “i ₃ ” as the refrigerant is further cooled by heat exchange with water (fluid to be heated), liquefied to a constant pressure inside the condenser 2 and condensed into saturated liquid (from B to C). And the enthalpy reaches “i ₄ ” when the refrigerant is in the supercooled state “D” at the outlet 7 of the condenser 2. The enthalpy change from “i ₃ ” to “i ₄ ” is much greater in FIG. 3b of the present invention compared to that of FIG. 5b, which indicates an increase in subcooling. In the process of throttle expansion from “D” to “E”, the refrigerant flows through the expansion valve (3) into the series condenser / evaporator (22) in an isenthalpy process of i ₄ = i ₅ ”, where the refrigerant evaporates completely. (E to F) and enter the “i ₆ ” state. Refrigerant with an enthalpy of “i ₆ ” is sucked into the compressor (1).

2a to 2c show a detailed view of one embodiment of the condenser used in the present heat pump system, wherein the condenser is a double tube consisting of a corrugated inner tube 32 and an outer tube 31 having a wire fin 33. 30 is formed. The coolant flows from the upper side through the gap between the inner tube 32 and the outer tube 31, and the fluid to be heated flows from the lower side through the inner tube 32.

In this way, according to the heat pump system of the present invention, high quality hot water or steam having an extremely high temperature of about 120 ° C. has a large temperature difference of 30 ° C. to 100 ° C., usually 50 ° to 90 ° C., between the condenser inlet and the outlet. Is discharged due to the temperature difference. Low temperature water of about 7 ° C. is made to be used alone or simultaneously with high temperature water.

Some embodiments of the invention are described.

Example 1

As shown in FIG. 1A, a two stage heat pump unit has a condenser having the structure shown in Table 1 below under the conditions shown in Table 2 below, water as a fluid and a fluid to be cooled when heated, and a hot cycle and a low temperature. It was operated using Flon R-114 and Flon R-22, respectively, as refrigerant for the side cycle.

TABLE 1

TABLE 2

Note: * Heat exchange rate = water flow rate × (water outlet temperature-water inlet temperature), * * supercooling = saturation temperature of refrigerant-outlet temperature of refrigerant.

The properties of the high boiling point refrigerant (Flon R-114) in the heat pump cycle according to the invention were measured and the obtained Moliere diagram values are shown in Table 3 below in comparison with the conventional heat pump cycle.

TABLE 3

It corresponds to the Moliere diagram of FIG. 5B. The following values are calculated from Table 3 above.

From Table 3, the enthalpy difference of the refrigerant liquid in subcooling is larger in the present invention than in the conventional heat pump operation.

Furthermore, the relationship between the subcooling and the coefficient of performance of the refrigerant in the condenser (Flon R-14) was examined and the results obtained are shown in Table 4 below.

The measurement conditions are as follows.

ㆍ Saturation Pressure 18.2kgf / cm ²

Saturation temperature (T _C ) 112.0 ℃

ㆍ Water condenser inlet temperature (t _w1 ) 19.1 ℃

ㆍ Enthalpy at the inlet of the compressor (i ₆ ) 145.4 kcal / kg

ㆍ Enthalpy at the outlet of the compressor (i ₁ ) 148.8 kcal / kg

TABLE 4

Hot water of about 99 ° C. was discharged at a temperature difference of about 80 ° C. at the outlet 7 of the condenser 2, and low temperature water of 7 ° C. was discharged to an inlet temperature of 12 ° C. at the outlet 19 of the evaporator 14. Ejected from

Example 2

The heat pump unit as shown in FIG. 1B is a refrigerant of dichlorodifluoromethane (Flon R-12), a condenser of the structure shown in Table 5 below, and a fluid to be heated and cooled under the conditions shown in Table 6 below. It was operated using water as.

The data are also shown in Table 6.

TABLE 5

TABLE 6

The temperature gradient and Moliere plot of this heat pump cycle are shown schematically in FIGS. 4a and 4b, respectively.

The characteristics of R-12 in a heat pump system showing the Moliere diagram (Fig. 4b) are shown in Table 7 compared to the conventional heat pump cycle.

TABLE 7

Note: The symbols “A” to “F” represent the states of FIG. 4b, and the symbols “a” to “f” represent the states of the Moliere diagram corresponding to FIG. 5b in the conventional heat pump system. .

In Table 7, the following performance values were calculated.

In this way, hot water of about 96 ° C was used as a temperature difference of about 76 ° C, and low temperature water of 7 ° C was used as a temperature difference of about 5 ° C.

As described so far, the present invention relates to a subcooling of a refrigerant in a condenser by heat exchange in a reverse flow manner with the fluid to be heated and by subcooling the refrigerant condensate with a subcooling of at least 20% of the temperature difference between the refrigerant saturation temperature and the condenser inlet temperature. It provides a heat pump system having a heat pump cycle in which the difference in enthalpy of the refrigerant liquid increases and the temperature difference between the inlet and outlet temperatures increases. Thus, the heat exchange amount, which is determined by the flow rate of the fluid and the temperature difference between the inlet and outlet temperatures of the fluid, is increased even with a small fluid flow rate. Therefore, the heat exchange efficiency as well as the outlet temperature of the fluid such as water can be much higher.

The high subcooling and increased enthalpy difference in the subcooling zone performed by the reverse flow process improves the coefficient of performance and makes it possible to instantaneously produce high quality water at 120 ° C. available. Furthermore, since the flow rate of the refrigerant can be reduced, the capacity of the compressor can be reduced, thereby reducing installation and operating costs.

Claims (12)

A heat pump system comprising a heat pump cycle of a refrigerant comprising a compressor, a condenser, an expansion valve, and an evaporator coupled in sequence, wherein the condenser further comprises a through passage for the fluid to be heated, the condenser between the refrigerant and the fluid. It consists of a double- or multi-tube heat exchanger of a completely reverse flow type, and when the system is operated, the refrigerant is liquefied and condensed while the refrigerant exchanges heat with the fluid in a reverse flow manner. Wherein the condenser is operated to cool to at least 20% subcooling of the temperature difference between the condenser inlet temperatures such that hot fluid is discharged from the outlet of the condenser. The heat pump system according to claim 1, wherein the evaporator further includes a through passage for the fluid to be cooled, so that the low temperature fluid or the high temperature fluid and the low temperature fluid are discharged. 3. The heat pump system according to claim 2, wherein the evaporator is configured as a heat exchanger in the form of completely reverse flow between the refrigerant and the fluid. 2. The heat pump system of claim 1, wherein the condenser is configured to discharge a hot fluid having an outlet temperature of 30 ° C to 100 ° C higher than the inlet temperature of the fluid. The heat pump system according to claim 1, wherein the condenser is configured to discharge a high temperature fluid having an outlet temperature of 50 ° C to 90 ° C higher than the inlet temperature of the fluid. 2. The heat pump system of claim 1, wherein the condenser is configured to operate at an outlet temperature of the fluid above a saturation temperature of the refrigerant. A high temperature side heat pump cycle of a high boiling point refrigerant comprising a compressor, a condenser and an expansion valve; A low temperature side heat pump cycle of a low boiling point refrigerant comprising a compressor, an expansion valve and an evaporator; A heat pump system comprising a series condenser / evaporator for coupling both cycles at a position between a compressor and an expansion valve of the high temperature side heat pump cycle and between an compressor and an expansion valve of the low temperature side heat pump cycle, wherein the condenser Further comprises a through passage for the fluid to be heated, wherein the condenser is comprised of a double or multi-pipe heat exchanger in full reverse flow between the high boiling point refrigerant and the fluid, and when the system is operated, the high boiling point refrigerant The condenser is operated such that after the high boiling point refrigerant is liquefied and condensed during the heat exchange with the fluid, the condenser is cooled to a subcool of at least 20% of the temperature difference between the saturation temperature of the high boiling point refrigerant and the condenser inlet temperature of the fluid. Heat pump system, characterized in that the hot fluid is discharged from the outlet of the condenser . 8. The heat pump system according to claim 7, wherein the evaporator further comprises a through passage for the fluid to be cooled, so that the hot fluid and the cold fluid are discharged simultaneously. 9. The evaporator of claim 8, wherein the evaporator consists of a heat exchanger in full reverse flow between the low boiling point refrigerant and the fluid to be cooled, and the series condenser / evaporator consists of a heat exchanger in full reverse flow between both refrigerants. Heat pump system. 8. The heat pump system of claim 7, wherein the condenser is configured to discharge a hot fluid having an outlet temperature of 30 ° C to 100 ° C higher than the inlet temperature 2 of the fluid. 8. The heat pump system of claim 7, wherein the condenser is configured to discharge a high temperature fluid having an outlet temperature of 50 ° C to 90 ° C higher than the inlet temperature of the fluid. 10. The heat pump system of claim 9, wherein the condenser is configured to discharge a hot fluid having an outlet temperature of 50 ° C to 90 ° C higher than the inlet temperature of the fluid.

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